Method and apparatus for controlling ignition timing in a compression-ignition engine operating in an auto-ignition mode

ABSTRACT

A method for controlling timing of ignition of a fuel charge in a compression-ignition engine operating in a controlled auto-ignition mode wherein the engine includes controllable intake and exhaust valve actuation systems is described. The method comprises determining a preferred ignition timing for a cylinder charge and a mass of the fuel charge based upon operator torque request. A portion of the fuel charge is partially oxidized during a negative valve overlap period immediately prior to a compression stroke. Magnitude of the portion of the fuel charge is based upon the preferred ignition timing of the cylinder charge. A remainder of the fuel charge is injected into the cylinder during the compression stroke.

TECHNICAL FIELD

This invention relates to operation and control of compression-ignitionengines operative in an auto-ignition mode.

BACKGROUND OF THE INVENTION

The statements in this section merely provide background informationrelated to the present disclosure and may not constitute prior art.

Compression-ignition engines, e.g., diesel engines, offer benefitsincluding improved fuel economy. Manufacturers of vehicles, includingpassenger cars, commercial trucks, construction, and agriculturaltractors, must meet stringent emissions requirements to market theirproducts.

Controlling ignition timing, i.e., retarding or advancing timing, on adiesel engine is a valuable control option, in order to achieve lowsmoke emissions. For example, it has been demonstrated that in a systemoperating in lean air/fuel ratio in a compression-ignition mode with apremixed cylinder charge (i.e., controlled auto-ignition orhomogeneous-charge compression-ignition, or HCCI), it is necessary tocomplete the injection of all fuel before ignition occurs, in order toachieve benefits related to lowered particulate matter (PM) emissions.

In an engine operating in auto-ignition mode, combustion of a cylindercharge is flameless, and spontaneously occurs throughout the entirecombustion chamber volume. The homogeneously mixed cylinder chargeauto-ignites as the cylinder charge is compressed and its temperatureincreases. The ignition timing of auto-ignited combustion depends oninitial cylinder charge conditions including, primarily, temperature,pressure, and composition of the cylinder charge. Thus, it is importantto coordinate engine control inputs, such as fuel mass, injectiontiming, and intake and exhaust valve motion, to ensure robustauto-ignition combustion.

Timing of fuel injection is currently used to control ignition andcombustion timing in diesel engines. Injection timing during the maincompression stroke is the main alternative control option forcontrolling ignition timing in diesel engines. By retarding timing ofinjection, ignition of the cylinder charge is retarded. However, it isnecessary to complete the injection of all fuel before ignition occurs.

It is well known that techniques for achieving low NOx emissions andparticulates associated with diesel auto-ignition combustion are limitedto moderate loads, because ignition of the cylinder charge is too rapidat high loads. It would be useful to have a method for retardingignition timing that is not related to injection timing. It would beuseful to expand the operating range of a compression-ignition engine inauto-ignition mode to improve emissions performance and fuel economy.

It is well known from studies of gasoline HCCI engines with negativevalve overlap (NVO) that fuel injected during the negative valve overlap(recompression) advances combustion timing. The reason given for this isthat the fuel injected during negative valve overlap is partiallyoxidized, or reformed, releasing some heat so that the temperature ofthe trapped residual is increased. The presence of higher trappedresiduals results in a higher temperature at intake valve closing,leading to earlier ignition in a gasoline engine operating in HCCI mode.

At moderate and high loads in a premixed diesel engine, the ignitiondelay is decreased due to high in-cylinder temperature. As a result,combustion timing is over-advanced, resulting in excessive engine noiseand possibly higher soot emissions. Premixed diesel combustion reliesheavily on recirculated exhaust gas (EGR) to retard combustion. Duringtransient operation of premixed diesel combustion, a delay in EGRreaching the cylinder can result in unstable ignition timing. At highload, engine noise becomes a problem because sufficient EGR cannot beinducted into the combustion chamber.

There is a need to control ignition timing independently of injectiontiming in a compression-ignition engine operating in an auto-ignitionmode. The benefits of so operating include expanding the dynamicoperating range of the engine in the auto-ignition mode, improvingemissions performance and fuel economy, and, minimizing engine noise.

SUMMARY OF THE INVENTION

In accordance with an embodiment of the invention, there is provided amethod and system for controlling timing of ignition of a cylindercharge in a compression-ignition engine including controllable intakeand exhaust valve actuation systems that is operating in a controlledauto-ignition mode. The method comprises determining a preferredignition timing for the cylinder charge and a fuel charge mass fordirect injection to the cylinder based upon operator torque request. Aportion of the fuel charge is partially oxidized by injection into thecylinder during a negative valve overlap period immediately prior to acompression stroke. Magnitude of the portion of the fuel charge is basedupon the preferred ignition timing of the cylinder charge. A remainderof the fuel charge is injected into the cylinder during the compressionstroke. These and other aspects of the invention are describedhereinafter with reference to the drawings and the description of theembodiments.

DESCRIPTION OF THE DRAWINGS

The invention may take physical form in certain parts and arrangement ofparts, the embodiments of which are described in detail and illustratedin the accompanying drawings which form a part hereof, and wherein:

FIG. 1 is a schematic drawing of an engine, in accordance with thepresent invention;

FIG. 2 is a data graph, in accordance with the present invention; and,

FIGS. 3A and 3B are data graphs, in accordance with the presentinvention.

DETAILED DESCRIPTION OF THE INVENTION

Referring now to the drawings, wherein the depictions are for thepurpose of illustrating the invention only and not for the purpose oflimiting the same, FIG. 1 comprises a schematic diagram depicting aninternal combustion engine 10, exhaust aftertreatment system 15, andcontrol module 5, constructed in accordance with an embodiment of thepresent invention.

The exemplary engine 10 comprises a multi-cylinder, direct-injection,compression-ignition, internal combustion engine having reciprocatingpistons 22 attached to a crankshaft 24 and movable in cylinders 20 whichdefine variable volume combustion chambers 34. The crankshaft 24 isoperably attached to a vehicle transmission and driveline to delivertractive torque thereto, in response to an operator torque request (‘TO_(—) REQ’). The engine preferably employs a four-stroke operationwherein each engine combustion cycle comprises 720 degrees of angularrotation of crankshaft 24 divided into four 180-degree stages ofintake-compression-expansion-exhaust, which are descriptive ofreciprocating movement of the piston 22 in the engine cylinder 20.

The engine preferably has an air/fuel operating regime that is primarilylean of stoichiometry. The skilled practitioner understands that aspectsof the invention are applicable to other engine configurations thatoperate primarily lean of stoichiometry, e.g., lean-burn spark-ignitionengines. During normal compression-ignition operation of thecompression-ignition engine, a combustion event occurs during eachengine cycle when a fuel charge is injected into the combustion chamberto form, with the intake air, the cylinder charge. The cylinder chargeis subsequently combusted by action of compression thereof during thecompression stroke. During the normal compression-ignition operation,the fuel charge is injected during the compression stroke. In-cylinderburned gases are generated due to incomplete combustion of the fuel andair mixture, which become exhaust gases when passed out of thecombustion chamber with opening of the engine exhaust valves during theexhaust stroke, which occurs after the expansion stroke. The exhaust gasis made up of regulated constituent elements preferably transformed bythe aftertreatment system to inert gases, typically includinghydrocarbons (‘HC’), carbon monoxide (‘CO’), nitrogen oxides (‘NO_(x)’),and particulate matter (‘PM’), among others. The engine includes sensingdevices to monitor engine operation, and actuators which control engineoperation. The sensing devices and actuators are signally or operativelyconnected to control module 5.

Sensing devices are installed on or near the engine 10 to monitorphysical characteristics and generate signals which are correlatable toengine and ambient parameters. The sensing devices preferably comprise acrank sensor 44 for monitoring crankshaft speed (RPM), a manifoldpressure sensor for monitoring manifold pressure (MAP) and ambientbarometric pressure (BARO), a mass air flow sensor for monitoring intakemass air flow (MAF) and intake air temperature (T_(IN)), and, an exhaustgas sensor 16 for monitoring states of one or more exhaust gasparameters, e.g., temperature, air/fuel ratio, and constituents (EXH).One skilled in the art understands that there may be one or more sensingdevices and methods for monitoring exhaust gas before, in the middle of,and after the exhaust aftertreatment system for purposes of control anddiagnostics. The operator input, in the form of the operator torquerequest (TO _(—) REQ) is typically obtained through a throttle pedal anda brake pedal, among other devices. The engine is preferably equippedwith other sensors (not shown) for monitoring operation and for purposesof system control. Each of the sensing devices is signally connected tothe control module 5 to provide signal information which is transformedby the control module to information representative of the respectivemonitored parameter. It is understood that this configuration isillustrative, not restrictive, including the various sensing devicesbeing replaceable with functionally equivalent devices and algorithmsand still fall within the scope of the invention.

The actuators are installed on the engine and controlled by the controlmodule 5 in response to operator inputs to achieve various performancegoals. Actuators include an electronically-controlled throttle devicewhich controls throttle opening to a commanded input (ETC), and aplurality of fuel injectors 12 for directly injecting fuel into each ofthe combustion chambers in response to a commanded input (INJ_PW), allof which are controlled in response to the operator torque request (TO_(—) REQ). There is an exhaust gas recirculation valve 26 and cooler(not shown), which controls flow of externally recirculated exhaust gasto the engine intake, in response to a control signal (EGR) from thecontrol module.

The fuel injector 12 is an element of a fuel injection system, whichcomprises a plurality of high-pressure fuel injector devices eachadapted to directly inject a fuel charge, comprising a mass of fuel,into one of the combustion chambers in response to the command signal,INJ_PW, from the control module. Each of the fuel injectors 12 aresupplied pressurized fuel from a fuel distribution system (not shown),and have operating characteristics including a minimum pulsewidth and anassociated minimum controllable fuel flow rate, and a maximum fuelflowrate. The minimum controllable fuel flow rate determines a lowerlimit for controllable fuel injection, including during a fuel reformingperiod that occurs during a negative valve overlap period, and, during amain injection event.

The engine 10 is equipped with a controllable valvetrain operative toadjust openings and closings of intake and exhaust valves of each of thecylinders, including any one or more of valve timing, phasing (i.e.,timing relative to crank angle and piston position), and magnitude oflift of valve openings. As depicted in FIG. 1, each cylinder includes anintake valve and an exhaust valve, the opening and closing of which arecontrolled by camshafts which are rotatably connected to the crankshaft.There is an intake air control device 38, comprising a variable camphaser (‘VCP’) which adjusts phasing of the opening and closing of theintake valve relative to the crankshaft rotation. The intake air controldevice 38 is preferably further mechanized to control valve lift of eachintake valve, referred to as variable lift control (‘VLC’). The variablelift system is operative to control lift of the intake valve to one oftwo or more distinct steps. The intake VCP/VLC system 38 controlsphasing, opening and closing times, and valve lift of the intake valve,in response to a control signal (VCP-INT) from the control module. Thereis an exhaust air control device 36, comprising a variable cam phaser(‘VCP’) which adjusts phasing of the opening and closing of the exhaustvalve relative to the crankshaft rotation. The exhaust air controldevice 36 is preferably further mechanized to control valve lift of eachexhaust valve, again referred to as variable lift control (‘VLC’). Thevariable lift system is operative to control lift of the exhaust valveto one of two or more distinct steps. The exhaust VCP/VLC system 36controls phasing, opening and closing times, and valve lift of theexhaust valve, in response to a control signal (VCP-EXH) from thecontrol module. The control module is operative to control phasings ofopenings and closings of the intake and exhaust valves to create the NVOperiod. This preferably includes executing control to simultaneouslyadvance the closing of the exhaust valves and retard the opening of theintake valve by substantially equal degrees of rotation, i.e., balancingthe phasings of the intake and exhaust valves. Similarly, the closing ofthe exhaust valves may be advanced and the openings of the intake valvesmay be retarded in a balanced manner.

The intake and exhaust air control devices may be actuated using one ofelectro-hydraulic, hydraulic, and electric control force. Other enginesystem components (not shown) may include an intake air compressingdevice, e.g., a variable geometry turbine device and air compressor, acharge air cooler, among others.

The engine operates un-throttled using diesel fuel or similar fuelblends in the auto-ignition combustion mode over a range of enginespeeds and loads which are typically determined during enginedevelopment and precalibrated into the control module. Conventionalcompression-ignition combustion is utilized at operating speed/loadconditions which are not conducive to operation in the auto-ignitioncombustion mode, and to obtain maximum engine power to meet the operatortorque request.

The control module 5 is preferably a general-purpose digital computergenerally comprising a microprocessor or central processing unit,storage mediums comprising non-volatile memory including read onlymemory (ROM) and electrically programmable read only memory (EPROM),random access memory (RAM), a high speed clock, analog to digital (A/D)and digital to analog (D/A) circuitry, and input/output circuitry anddevices (I/O) and appropriate signal conditioning and buffer circuitry.The control module has a set of control algorithms, comprising residentprogram instructions and calibrations stored in the non-volatile memoryand executed to provide the respective functions of each computer. Thealgorithms are typically executed during preset loop cycles such thateach algorithm is executed at least once each loop cycle. Algorithms areexecuted by the central processing unit and are operable to monitorinputs from the aforementioned sensing devices and execute control anddiagnostic routines to control operation of the actuators, using presetcalibrations. Loop cycles are typically executed at regular intervals,for example each 3.125, 6.25, 12.5, 25 and 100 milliseconds duringongoing engine and vehicle operation. Alternatively, algorithms may beexecuted in response to occurrence of an event.

The control module 5 executes algorithmic code stored therein to controlthe aforementioned actuators to control engine operation, includingthrottle position, fuel injection mass and timing, intake and/or exhaustvalve timing, phasing, and lift, and EGR valve position to control flowof recirculated exhaust gases. The control module is adapted to receiveinput signals from the operator (e.g., a throttle pedal position and abrake pedal position) to determine the operator torque request (T_(O)_(—) _(REQ)) and from the sensors indicating the engine speed (RPM) andintake air temperature (T_(IN)), and coolant temperature and otherambient conditions. The control module 5 determines, from lookup tablesin memory, instantaneous control settings for EGR valve position, intakeand exhaust valve timing and/or lift set points, and fuel injection massand timing

The invention comprises a method and control scheme for controllingtiming of ignition of the cylinder charge in the exemplarycompression-ignition engine described above, operating in auto-ignitionmode, thus increasing the dynamic speed/load operating range of theengine in the auto-ignition mode, with accompanying benefits accruingthereto.

The method comprises controlling timing of ignition of the cylindercharge while operating the engine in the controlled auto-ignition mode,including splitting injection of the fuel charge into two or moreinjection events during an engine cycle. The engine control includesexhaust recompression, wherein timing and phasing of the intake andexhaust valves are controlled to achieve the negative valve overlap(‘NVO’) period during which partial oxidation of fuel injected duringthis NVO period can take place. This partial oxidation is also sometimesreferred to as fuel reforming. A preferred ignition timing for thecylinder charge and a total mass of the fuel charge for direct injectionto the cylinder to meet the operator torque request are determined bythe control module, based upon known calibrations and engine operatingcharacteristics. A first portion of the fuel charge is injected duringthe NVO period immediately prior to the compression stroke, duringongoing engine operation. Mass of the first portion of the fuel chargeis based upon the preferred ignition timing of the cylinder charge. Aremaining portion of the fuel charge is injected into the cylinderduring the subsequent compression stroke. The mass of the first portionof the injected fuel charge is determined based upon the minimumcontrollable fuel flow rate of the injector. This mass is either zeropercent of the fuel charge, or can range between the minimumcontrollable fuel flow rate of the injector and approximately fiftypercent of the fuel charge. The remainder of the total mass of fuel isinjected during the subsequent compression stroke in accordance with theauto-ignition combustion process.

Referring now to FIG. 2, and Table 1, below, results of simulated engineoperation are described for the engine operated as described to controlignition timing by reforming fuel during the NVO period. The concept ofthe method is based upon a change in composition of trapped residual gasin the combustion chamber when the fuel injected during the NVO period,due to partial oxidation of the fuel. With diesel fuel, the reformingfuel produces a composition that is more difficult to ignite than dieselfuel itself. The simulated engine operation comprised full-cycle enginemodeling calculations performed with n-heptane, a diesel fuel surrogate,wherein 1.05 milligrams (mg) of fuel were injected before top deadcenter (TDC) of the NVO period and 13.95 mg fuel were injected duringthe compression stroke shortly before TDC of the main combustion stroke.The results demonstrate that some of the n-heptane injected before TDCduring the NVO period was partially oxidized, or reformed, to formmolecules such as methane, ethene, ethane, propene, 1-pentene, carbonmonoxide, and heptene. The molecules were trapped in the combustionchamber, and were mixed with intake air during the subsequent intakestroke, to produce a mixture of these molecules with n-heptane and airat intake valve closing. The molecules, both separately and whencombined together, are more difficult to auto-ignite than n-heptaneitself. The mixture of n-heptane and the molecules resulted in aretarded ignition timing compared to the case where 1.05 mg n-heptanewas injected after TDC of the NVO period and 13.95 mg fuel was injectedwith the same injection timing shortly before TDC of the main combustionstroke. In the latter case little or no partial oxidation of theinjected n-heptane occurred. Diesel fuel typically behaves similarly ton-heptane, because diesel fuel contains many long, straight-chainhydrocarbons that have very similar combustion characteristics ton-heptane. As a result of this similar structure to many components ofdiesel fuel, n-heptane has a high cetane number, similar tocommonly-used diesel fuels. The simulation results are depicted withreference to FIG. 2 and detailed in Table 1. The single injection eventhas an end of fuel injection (EOI2) occurring at 20 crank angle degreesbefore TDC; the dual injection event without fuel reforming has a firstinjection event of 1.05 mg fuel and an end of the first fuel injection(EOI1) occurring at 405.7 crank angle degrees after the previousTDC-compression, i.e., after TDC-intake and an end of second fuelinjection (EOI2) occurring at 20 crank angle degrees before TDC; and,the dual injection event with fuel reforming has a first injection eventof 1.05 mg fuel and an end of the first fuel injection (EOI1) occurringat 310.7 crank angle degrees after the previous TDC-compression, i.e.,before TDC-intake and an end of second fuel injection (EOI2) occurringat 20 crank angle degrees before TDC.

TABLE 1 Single Parameter Injection No Reforming Reforming Injected fuel,mg 15 15 Injection duration, 10 10 10 CA deg. EOI1, aTDC — 405.7 310.7combustion EOI2, aTDC −20 −20 −20 Injection amount 1, 0 1.05 1.05 mgInjection amount 2, 15 13.95 13.95 mg NMEP, bar 3.46 3.44 3.46 CA10,deg. aTDC 0.9 −0.35 3.8 CA50, deg. aTDC 3.0 1.9 6.4 Noise, MW/m² 4.1 3.43.5 T @ IVC, K 426 424 431

As described and depicted with reference to FIG. 2 and Table 1, at theconditions of the calculations the heat released during the partialoxidation of the n-heptane during the NVO period results in a highertemperature at intake valve closing, but a retarded combustion phasingcompared to the absence of this partial oxidation. A higher temperatureat intake valve closing tends to advance the auto-ignition timing. Thepartial oxidation products of the fuel, however, tend to retard theauto-ignition timing. This means that there is a competition between thetemperature effect and the composition effect. Under some conditions oneof these may dominate, and at other conditions the other may dominate.Referring to FIG. 3, the composition effect dominates for amounts offuel, injected before TDC of the NVO period, ranging from about 1 mg to4 mg, with the remainder of the 15 mg fuel charge injected during thecompression stroke.

Engine load and speed conditions under which the control modulepreferably executes the control scheme described herein comprise a loadrange from approximately 200 to 700 kPa brake-mean-effective-pressure(BMEP) and from approximately 1000 to 4000 RPM.

Referring now to FIG. 3, results of the simulated engine operation aredepicted for the engine operated as described to control ignition timingby reforming fuel during NVO, with varying amounts of reforming fuel.The results in FIG. 3A depict calculated CA10, i.e., a crank anglelocation at which 10% of the fuel mass is burned as a function of themass of reforming fuel (mg), for single injection operation, dualinjection operation with the end of injection (EOI) occurring beforeTDC, and dual injection operation with the end of injection (EOI)occurring after TDC. The results in FIG. 3B depict calculated CA50,i.e., a crank angle location at which 50% of the fuel mass is burned asa function of the mass of reforming fuel (mg), for single injectionoperation, dual injection operation with the end of injection (EOI)occurring before TDC, and dual injection operation with the end ofinjection (EOI) occurring after TDC. In each case of dual injection, theremaining fuel was injected such that the end of injection occurred 20crank angle degrees bTDC of the subsequent injection. These resultsdemonstrate that ignition of the cylinder charge is retarded with thedual injection operation having the end of injection (EOI) occurringbefore TDC of the NVO period. Thus, injecting some of the fuel duringthe NVO period retards the ignition timing, and reduces the enginenoise, allowing an expansion of the controlled auto-ignition operatingrange to higher load conditions.

While the invention has been described by reference to certain preferredembodiments, it should be understood that changes can be made within thespirit and scope of the inventive concepts described. Accordingly, it isintended that the invention not be limited to the disclosed embodiments,but that it have the full scope permitted by the language of thefollowing claims.

1. Method for controlling timing of ignition of a cylinder charge in acompression-ignition engine operating in a controlled auto-ignitionmode, the compression-ignition engine including controllable intake andexhaust valve actuation systems, the method comprising: determining afuel charge for injecting diesel fuel into the cylinder to create acylinder charge based upon an operator torque request; determining apreferred ignition timing for the cylinder charge; partially oxidizing aportion of the fuel charge in the cylinder, wherein mass of thepartially oxidized portion of the fuel charge is based upon thepreferred ignition timing of the cylinder charge; and, injecting aremainder of the fuel charge into the cylinder during an immediatelysubsequent compression stroke.
 2. The method of claim 1, whereinpartially oxidizing the portion of the fuel charge in the cylindercomprises injecting the portion of the fuel charge into the cylinderduring a negative valve overlap period immediately prior to thecompression stroke.
 3. The method of claim 2, further comprisinginjecting the portion of the fuel charge into the cylinder during thenegative valve overlap period such that an end of the injecting occursprior to a top-dead-center of piston travel during the negative valveoverlap period.
 4. The method of claim 3, further comprising the end ofthe injecting occurring twenty degrees prior to a top-dead-center ofpiston travel during the negative valve overlap period.
 5. The method ofclaim 2, wherein the negative valve overlap period comprises: adjustingopenings and closings of intake and exhaust valves to create a negativevalve overlap period between a closing of the exhaust valve and anopening of the intake valve for the cylinder.
 6. The method of claim 2,further comprising: increasing the portion of the fuel charge injectedduring the negative valve overlap period to increasingly retard theignition timing.
 7. Method for retarding ignition timing in acompression-ignition engine operating in a controlled auto-ignitionmode, the compression-ignition engine including controllable intake andexhaust valve actuation systems, the method comprising: determining afuel charge for direct injection of diesel fuel into a cylinder to meetan operator torque request; adjusting openings and closings of theintake and exhaust valves to create a negative valve overlap periodbetween closing of the exhaust valve and opening of the intake valve forthe cylinder; injecting a portion of the fuel charge into the cylinderduring the negative valve overlap period wherein an end of the injectingoccurs prior to a top-dead-center point of piston travel; and, injectinga remainder of the fuel charge into the cylinder during an immediatelysubsequent compression stroke.
 8. The method of claim 7, whereinadjusting opening and closing of the intake and exhaust valves to createa negative valve overlap period between closing of the exhaust valve andopening of the intake valve for the cylinder comprises adjustingphasings of exhaust and intake camshafts operative to actuate theexhaust valve and the intake valve.
 9. The method of claim 8, whereinadjusting opening and closing of the intake and exhaust valves to createa negative valve overlap period between closing of the exhaust valve andopening of the intake valve for the cylinder further comprisescontrolling magnitude of lift of the exhaust valve and the intake valvewhen the valves are open.
 10. The method of claim 8, wherein adjustingopenings and closings of the intake and exhaust valves to create thenegative valve overlap period comprises adjusting the intake and exhaustvalve actuation systems to advance closing time of the exhaust valve andretard opening time of the intake valve by substantially equal amounts.11. The method of claim 8, wherein injecting the portion of the fuelcharge into the cylinder during the negative valve overlap periodfurther comprises controlling the portion of the fuel charge based upona preferred ignition timing of the cylinder.
 12. The method of claim 11,further comprising: increasing the portion of fuel injected during thenegative valve overlap period to increasingly retard the ignitiontiming.
 13. The method of claim 7, wherein injecting the portion of thefuel charge into each combustion chamber during the negative valveoverlap period further comprises starting the injecting of the portionof the fuel charge before top dead center of the negative valve overlapperiod.
 14. The method of claim 13, further comprising injecting theportion of the fuel charge into the cylinder during the negative valveoverlap period such that an end of the injecting occurs prior to thetop-dead-center of piston travel.
 15. The method of claim 14, furthercomprising the end of the injecting occurring twenty degrees prior tothe top-dead-center of piston travel.
 16. Article of manufacture,comprising a storage medium containing a machine-executable programoperative to control timing of ignition of a cylinder charge in acompression-ignition engine selectively operating in a controlledauto-ignition mode, the compression-ignition engine includingcontrollable intake and exhaust valve actuation systems, the programcomprising: code to determine a total mass of a diesel fuel charge fordirect injection to the cylinder to create a cylinder charge based uponoperator torque request; code to determine a preferred ignition timingfor the cylinder charge; code to adjust openings and closings of theintake and exhaust valves to create a negative valve overlap periodbetween closing of the exhaust valve and opening of the intake valve forthe cylinder; code to control fuel injection to inject a portion of thediesel fuel charge into the cylinder during the negative valve overlapperiod wherein an end of the injecting occurs prior to a top-dead-centerpoint of piston travel; and, code to control fuel injection to inject aremainder of the diesel fuel charge into the cylinder during animmediately subsequent compression stroke.
 17. The article ofmanufacture of claim 16, wherein the code to control fuel injection toinject a portion of the diesel fuel charge into the cylinder during thenegative valve overlap period further comprises: code to control theportion of diesel fuel injected during the negative valve overlap periodbased upon the preferred ignition timing for the cylinder charge. 18.The article of manufacture of claim 17, wherein the code to control theportion of diesel fuel injected during the negative valve overlap periodbased upon the preferred ignition timing for the cylinder chargecomprises code to increase the injected portion of the diesel fuelcharge into the cylinder during the negative valve overlap period toincreasingly retard the ignition timing.
 19. The article of manufactureof claim 18, wherein the portion of diesel fuel injected during thenegative valve overlap period ranges between a minimum controllableamount and approximately fifty percent of the diesel fuel charge.